Non-contacting mechanical face seal of the gap-type

ABSTRACT

A rotary mechanical face seal for preventing leakage of a fluid, either liquid or gas, across two relatively rotating surfaces. The seal is of the non-contacting mechanical type in which a predetermined spacing is maintained between the relatively rotatable sealing washers by formation of a pressurized fluid film between the relatively rotating surfaces. The opposed faces are constructed to provide a separation of the faces at startup and thereby prevent damage or destruction to the faces before design pressures are achieved. This is achieved by providing at least one of the opposed faces with a deflection bearing portion which causes the necessary deflections under load. The bearing portion comprises a spaced series of lift pad sections. A variety of lift pad constructions can be used. The seal also includes a hydrodynamic radial bearing which centers the shaft in the static state to avoid shaft runout.

This application is a continuation of application Ser. No. 07/785,005filed Oct. 30, 1991, now U.S. Pat. No. 5,246,295.

FIELD OF THE INVENTION

This invention relates to a rotary mechanical gap type end face seal forsealing the space between a rotating shaft and its housing.

BACKGROUND OF THE INVENTION

Rotary mechanical gap type radial face seals are used for effecting aseal between relatively rotatable members such as a shaft and a housingby controlling leakage of the fluid from one region of high pressure toa second region of lower pressure.

Such gap type seals are typically formed with two sealing members. Oneof the members is fixed so that no movement occurs axially relative tothe shaft, and this is referred to as a fixed seal member or, in thiscase, mating ring. The other element is movable axially along the shaftand is sometimes referred to as a floating element or in this particularcase, the sealing ring. The elements each include sealing faces whichare located in opposed relationship to each other. The sealing faces aresuch that, in response to fluid pressure, spring pressure, or both, asealing relationship will be obtained between them to prevent leakageout along the shaft.

During operation the relatively rotatable sealing members are kept fromtouching one another during the operation of the seal. Thischaracteristic makes them ideal for very high speeds, since there is noappreciable wear of the sealing members and hence no appreciabledestructive heat produced by their relative rotation.

To keep the sealing members from touching one another under designoperating conditions, a fluid pressure is created between theconfronting seal faces. When this pressure exceeds the pressure(typically called spring pressure) tending to bring the seal facestogether the seal faces are separated. The degree of separation iscontrolled by the action of the fluid as it passes between the faces tothe low pressure side of the seal. At startup of the equipment in whichthe seal is installed, the design fluid pressure is not available.Accordingly, the seal faces are in contact while the pressure in theseal chamber is building up to the design pressure. Such contact, evenif brief, may be sufficient to create a heat and wear condition at thefaces sufficient to destroy the seal.

One known gap seal is disclosed in James F. Gardner U.S. Pat. Nos.3,499,653 and 3,804,424. The seal disclosed in these patents is amechanical end face seal which operates with a gap between the opposedradial sealing faces of the sealing rings to permit controlled leakage.These rings have flat, radially extending surfaces which sealinglyengage one another. Shallow spiral grooves are formed in the outerperiphery of one of the relatively rotatable seal members, preferablythe stationary one, to create a pump, the direction of the spiral beingsuch that the fluid to be sealed is forced between the seal members toseparate and lubricate them at start-up. The spiral grooves, however,are effective in only one direction of relative rotation so that theseal is directional and may be objected to for that reason.

A similar known arrangement uses T-shaped grooves instead of spiralgrooves. The function is quite similar. The grooves generate pressure toforce the faces axially apart.

There are problems with known gap seal designs. For instance, thedimensions of these recesses are critical and difficult to manufacturebecause of the tight tolerances that are required. Also, any contaminantin the fluid has a significant detrimental effect on the performance ofthe seal. Accordingly, the hydrodynamic performance range is limitedbecause of the fixed geometry in this structure.

The environment of the present invention is similar to that of thenknown constructions. Accordingly, reference will be had to one knownconstruction.

Those skilled in the art will recognize the environment of gap sealsused today is typically far more sophisticated than the environmentdescribed below. The following description is, however, intended simplyto illustrate an example of the type of seal improved by the presentinvention.

FIG. 1 shows a gap seal having a rotor or mating ring 10 mounted on ashaft 11 and having a substantially radially disposed sealing surface 12which has been appropriately lapped to be perfectly flat and smooth. Themating ring 10 is preferably made in the form of a washer which isfinished and lapped independently of shaft 11 and is then assembled withrespect to said shaft, in a manner so as to be rotatable therewith.Adjacent mating ring 10 and surrounding shaft 11 is a sealing washer ormember 13 having a sealing surface 14 adjacent to and confrontingsealing surface 12 on the mating ring 10. The sealing washer issometimes called the sealing ring; that parlance will be adoptedhereinafter. Sealing ring 13 is formed with an axially extending sleeve15 which fits into an appropriate opening 16 in a housing 17 throughwhich shaft 11 extends. The opening 16 is enlarged at 18 to form a sealchamber in which the mating ring 10 and sealing ring 13 may operate.

Seal chamber 18 is filled with a fluid, either a gas or a liquid, as thecase may be, at any desired pressure above atmospheric. The pressure inopening 16, on the other hand, may be atmospheric pressure so that thefluid in seal chamber 18 tends to move radially inwardly betweensurfaces 12 and 14 and into the space between the sleeve 15 and shaft11, to the opening 16. A seal of any suitable character, such as anO-ring 19, supports sleeve 15 and sealing ring 13 resiliently in opening16 to allow said sealing ring to move axially in opening 16, as well asradially, to a limited extent. A very light spring 20 may be retainedbetween sealing ring 13 and the radial wall 21 of chamber 18 to urgesealing ring 13 against mating ring 10 when there is no pressure inchamber 18.

Sometimes the sealing surface 14 is made slightly convex by a lappingoperation to provide a wedge-shaped space at the radially outer regionsof sealing ring 13 to initiate and maintain the separation of thesurfaces 12 and 14 under operating conditions. When this is done, theactual separation at the low pressure side of surface 14 is very smallso that the separation shown in FIG. 1 is greatly exaggerated forpurposes of illustration. The curvature of the surface 14 is likewiseconsiderably exaggerated.

In theory, under operating pressures, sealing ring 13 will be pushedaway from surface 12 a predetermined distance and will then maintainthat distance or separation regardless of axial or radial movements ofthe mating ring 10, the sealing ring 13 being compelled to follow suchmovements by the pressure effect of the fluid being sealed. This actionis such that should any external forces be present tending to reduce thegap between surfaces 12 and 14, the forces of the fluid upon the movablesealing ring 13 will counter such external force and move sealing ring13 to the right, as shown in FIG. 1, until the designed gap is created.Similarly, should the external forces be such as to tend to increase theopening between surfaces 12 and 14 above the designed opening or gap,the said forces of the fluid will urge sealing ring 13 to the left, asviewed in FIG. 1, to reduce the gap to the designed size.

Whenever the pressure of the fluid is below that for which the seal isdesigned to operate as a gap seal, and the sealing members are rotatingrelative to one another, sealing ring 13 will contact mating ring 10 andthereby establish frictional contact between surfaces 12 and 14. Thiscontact is augmented by spring 20, the function of which is to close thegap between surfaces 12 and 14 when the equipment is not operating andthereby prevent a leakage of the fluid along shaft 12 into opening 16and also to prevent dirt particles and other harmful substances fromgetting between the seal surfaces 12 and 14. Although such contact isdesirable when there is no relative rotation between the mating ring 10and sealing ring 13, it is however, highly undesirable as the relativespeeds and pressures between surfaces 12 and 14 increase to the designedspeeds and pressures, since even during the brief period that theequipment is getting up to speed or slowing down to stop, sufficientfriction and heat can be generated to destroy the surfaces 12 and 14,particularly if the fluid sealed has low lubricating qualities such as agas.

Undesirable friction and heat are eliminated in conventional gap sealsby providing shallow spiral grooves in one of the surfaces 12 or 14. Theshape of the grooves is such as to cause fluid in chamber 18 to beforced radially inwardly even at relatively slow speeds of rotation ofrotor 16, across the inner regions of surfaces 12 and 14. A hydrodynamicwedge is thus created which provides sufficient pressure to separate thesurfaces 12 and 14 and forms a film of the fluid being sealed on whichthe surface 12 rides. This, in turn, eliminates or prevents, any directcontact between surfaces 12 and 14 and prevents the generation ofdestructive friction and heat.

Referring to FIG. 2, the spiral grooves are shown at 22. The preciseshape and size of the grooves depends largely upon degree ofeffectiveness required of them. In the form shown in FIG. 2, they extendspirally inward across slightly more than one-half the surface 14 andterminate at 24. They should not of course, extend across the entiresurface 14 since they would then provide a leak path across the seal.The area of the grooves illustrated is a little less than one-fourth thearea of said surface. The area between the grooves is indicated at 23.The groove depth, area, helix angle and the distance at which thegrooves terminate may be varied to suit different operating conditions.The depth of grooves 22 is preferably two or three times the actualminimum clearance or gap between surfaces 12 and 14 when the seal is inoperation.

The shallow grooves 22 may be formed in any known way. Etching is themost typical.

Since the grooves 22 are spiraled, the relative direction of rotationbetween the surfaces 12 and 14 must be such as to cause the fluid to beforced radially inwardly through the grooves 22. This means that thesurface 12 must rotate in the same direction as the direction in whichthe grooves 22 are spiraled. This, in turn, limits the use of the sealto an installation in which the shaft is rotating in the direction forwhich the seal is designed. Such limitation, however, can be eliminatedby the known constructions, which employ two or more sealing rings or byusing a symmetrical groove formation as discussed below.

FIG. 3 shows an alternative groove formation known in the art. Morespecifically, FIG. 3 shows the surface of a ring having a series ofT-shaped grooves 122 formed therein. The T-shaped grooves aresymmetrically disposed across the surface of the ring. The grooves 122function in essentially the same manner as the grooves 22 shown in FIG.2. Specifically, grooves function to cause hydrodynamic pumping effectsso as to cause separation of the opposed sealing faces. As with thespiral grooves shown in FIG. 2, the T-shaped groove should not extendacross the entire ring surface. One advantage of the grooveconfiguration shown in FIG. 3 is that it is symmetrical so that itoperates in the same way regardless of the direction of rotation. Thus,this type of ring formation can be used for bi-directional sealing. Likethe grooves shown in FIG. 2, the T-shaped grooves shown in FIG. 3 areextremely shallow and typically formed by etching or some otherrelatively sophisticated.

Another known design is shown in FIG. 4. In accordance with this design,a circumferentially spaced series of tapered lands 222 are formed alongthe outer periphery of one of the sealing ring and the mating ring. Theland is tapered such that it gradually recedes from the surface. At themore recessed end, a step-down is formed to form a sharply recessedportion 223. Thus, with reference to FIG. 4, the land tapers from theleft downward toward the right with a drop off at the recess 223.Because of the non-symmetrical nature of this tapered land, this type ofgroove formation is not suitable for hi-directional operation. Again,however, bi-directional operation can be provided by using two similarrings as is known in the art.

Typically tapered lands of the type shown in FIG. 4 must be provided byprecision machining on the smooth face of either the sealing ring or themating ring. It is easy to appreciate that such precise machining isdifficult and expensive. In operation the ring formation shown in FIG. 4operates in essentially the same way as the ring configuration shown inFIG. 2 and FIG. 3. In particular, the surface formation causes apressurization of the fluid between the sealing faces causing a radialgap to form between the sealing faces.

The addition of spiral, T-shaped grooves or tapered lands provideshydrodynamic load support for the sealing ring 13. Upon the start ofrotation, fluid is pumped between the faces of the seal, and at a givenRPM, the hydrodynamic load support becomes sufficient to give completeseparation. The seal is operable at zero pressure because of the springforce pushing the surfaces together.

It is understood that the grooves may be formed in surface 12 of matingring 10 in FIG. 1, instead of in the confronting surface on the sealingring 13. It is also understood that the curvature, if desired, may beformed on the surface or surfaces of the mating ring with the groovedsealing ring having a flat surface. It is also possible to eliminate thecurvature especially if, as with the present invention, another way ofachieving the desired effect is provided.

The present invention is intended to replace known designs in which thesurface of the mating ring is etched to create surface grooves whichcreate gas dynamic effects. In these known designs, the groove is veryshallow--on the order of millionths of an inch deep. Typically, thegrooves are formed by a photoetch process which is complicated andexpensive. Moreover, even with these extremely shallow grooves, there isa step at the transition between the groove and the surface in which thegroove is formed. This step tends to create non-laminar flow of thesealing fluid. It is known that the best sealing effects are achievedwhen the laminar flow of the sealing fluid is maintained.

There remains a need in the art to have a controllable mechanical sealwhere the thickness of the lubricating fluid film can be maintained at apractical thickness and one in which the manufacturing tolerances arereduced. It is further desirable to have a seal arrangement wherecontaminants are less likely to impact upon the performance of the seal,and, one in which the seal can self adjust for any shaft misalignment.Further, it is desirable to have a seal which will operate over abroader range and reduce any ultimate seal wear by obtaining an optimumfilm thickness over a wide range of operating conditions.

One improved mechanical face seal is disclosed in the present inventor'sprevious U.S. Pat. No. 4,738,453. In that patent, a controllablemechanical seal was disclosed for a machine having a housing and a shaftthat is rotatable relative to the housing. The seal includes astationary cylindrical seat member and a rotatable cylindrical nosepiece. The nose piece is fitted with a plurality of lift pads that areheld by the nose piece and the nose piece is also fitted with a fluiddam which defines a radial face surface. One or the other of the parts,either the seat, that is, or the nose piece will be rotatable with theshaft and suitable means will bias one element toward the other. Thelift pads are particularly formed as stool like units, having flexibleleg ligaments that extend at an angle to the pad face, so that the padface may move in up to three degrees of freedom to form a fluid filmbetween the pad face and the seat member to adjust for shaftmisalignment and to provide equal loading among the lift pads. Equalloading among pads in the longitudinal shaft axis direction is providedby dog leg type bends in the ligament construction.

In any mechanical face seal, it is important that the two rings, themating ring and the sealing ring are aligned such that their faces arein flush contact. Often, this is done by making one of the ringsfloating and spring biasing the two rings together so that the two ringsare pressed into flush contact by the springs.

Many known designs employ complicated expensive alignment, mechanisms orarrangements. There is still a need for a simple inexpensive way toalign the sealing faces relative to one another. There is also a needfor a more simple spring biasing construction.

In any gap seal the sealing ring and the mating ring must rotate aboutconcentric axes to ensure proper performance. Since the mating ringtypically rotates with the shaft and the sealing ring is secured in thehousing, the shaft must be supported for rotation about a fixed axis inthe housing in order for the gap seal assembly to function properly. Inother words, the eccentricity of the shaft must be minimized to avoidshaft runout.

In the past this has meant that ball bearings must be used to supportthe shaft for rotation about a fixed axis. Ball bearings are far fromideal. They tend to wear rapidly at high speeds and are expensive forthat reason. Conventional hydrodynamic bearings cannot be used, however,because the shaft position is not fixed until the shaft reaches designspeed. The eccentricity of the shaft during the start-up would lead toundesirable movement of the sealing faces which would defeat the gapseal. There remains a need for a durable inexpensive bearing forsupporting the shaft in a gap seal assembly for rotation about a fixedaxis.

SUMMARY OF THE INVENTION

The present invention is directed to an improved mechanical face sealcapable of achieving improved hydrodynamic sealing effects with aconstruction which is easier to manufacture than conventional etched ortapered land mechanical seal constructions. The gap-type rotarymechanical seal assembly includes a housing, a shaft, and relativelyrotatable sealing members having substantially radially disposedconfronting sealing faces. One of the sealing members is secured to theshaft for rotation therewith and the other sealing member is supportedin the housing. One of the sealing members is axially movable relativeto the other member. At least one of the sealing members includes a ringhaving a bearing portion and a sealing portion. The bearing portionincludes a plurality of spaced pad sections. Each pad section has ashape and support structure such that under load the sealing surface ofthe pad deflects to form a hydrodynamic wedge. At least one of thesealing member is supported on a support structure for movement with sixdegrees of freedom.

The assembly also includes a hydrodynamic bearing for supporting theshaft in the housing. The hydrodynamic bearing includes a plurality ofspaced bearing pads. Each bearing pad is supported by a supportstructure. At least one of the bearing pads is supported such that understatic load the bearing pad deflects such that the trailing edgecontacts the shaft so as to maintain the shaft in a centered position.

One aspect of the present invention is the provision of an improvedmating ring or seal ring. The improved ring is designed by treating thering as if it includes both a continuous smooth sealing portion and abearing portion. The bearing portion is constructed in accordance withthe deflecting pad bearing principles developed by the present inventorso as to cause a hydrodynamic effect under load which causes separationof the sealing ring and mating ring to achieve the necessary mechanicalseal. Generally, the sealing section of the ring is smooth and has asolid support. The bearing section, on the other hand, is designed todeflect under load to create the necessary hydrodynamic effect becauseof the slight separation between the sealing ring and the mating ringnecessary for proper mechanical hydrodynamic sealing. In some cases, theimproved ring is formed by undercutting the backside of the mating ringor sealing ring to cause deflection under load at the sealing face tocreate hydrodynamic pressurization of the fluid. The present inventorhas discovered that appropriate depressions or pockets on the sealingface under load can be provided by modifying the backside of either thesealing ring or the mating ring or by providing lift pads with a shapeor support structure. These modifications are on a much larger scalethan the microfine surface formations necessary in accordance with knowndesigns. In this way, the manufacturing cost of the seal is dramaticallyreduced because there is no need to be as precise. The seal may beconstructed either bidirectionally or unidirectionally.

As noted above, one alternative is to provide the bearing section withthe necessary support by simply undercutting a continuous surface toachieve the desired deflections under load. The bearing section may alsobe divided into separate segments each of which is connected to thecontinuous sealing section by a thin ligament. The necessary deflectionto obtain a hydrodynamic face is obtained by designing the pad shape orsupport structure in connection with the teachings of the presentinventor's previous applications. In this regard, applicant incorporatesby reference the teachings of his prior application Ser. No. 07/685,148filed Apr. 15, 1991 entitled "Hydrodynamic Bearings Having SpacedBearing Pads and Methods of Making Same".

In accordance with the present invention, the necessary recesses on thesealing face can be provided by undercutting the sealing face ratherthan forming recesses directly on the sealing face. Finite elementanalysis has shown that the portions of the sealing face which areundercut tend to deflect under load to form shallow recesses or pockets.Through proper location of the undercuts the recesses formed bydeflection can operate in the same manner as the recesses formeddirectly on the sealing face to provide a hydrodynamic sealing effectknown in the prior art. The precise location for any specificapplication is determined by a finite element analysis.

In accordance with the present invention, the hydrodynamic wedge isformed by deflection. As a result, the recesses tend to have a graduallytapered shape so that laminar flow across the sealing face ismaintained. This enhances the hydrodynamic sealing effect.

Another, advantage of the present invention is that the necessaryundercuts or pad shape or support structure are relatively easy tomanufacture into the sealing ring. In fact, in accordance with oneimportant aspect of the invention, the sealing or mating ring may be inthe form of a thin piece of plastic which can be manufactured byinjection molding or some other low cost process which is much lessexpensive than the photoetching or precision machining required in knowndesigns in which the recesses are formed directly on the sealing face.The plastic disk can be on the order of 0.1 inches thick. For example,with a plastic disk of about 0.1 inches thick a groove formed in thebackside to a depth of about 0.03 inches thick yields deflections ofabout 0.0002 inches deep on the front side of the plastic disk whenloaded in the case of an oil type seal. The deflection will be smallerin a gas type seal and therefore will require deeper undercuts. Ofcourse, other materials can be used for the disks, such as bronze,silicon carbide and the like.

In accordance with another aspect of the present invention, one of therings may be supported on a beam like support so that the ring canadjust its orientation so as to move into flush contact with the otherring. In this way, the ring has a self-aligning capability. Moreover,the beam-like support can be constructed to provide the necessary springforce to bias the sealing ring and mating ring into contact.

Another aspect of the present invention is the provision of ahydrodynamic bearing which can be used in a mechanical gap seal. Thisbearing, unlike conventional hydrodynamic bearings, is designed to holdthe shaft in position during start-up so that the mating ring and sealring always rotate about concentric axis. This is done by usingdeflection pad bearings of the type described in the applicant'sprevious application Ser. No. 07/685,148 filed Apr. 15, 1991 andincorporated above by reference. These deflection pad bearings are of aspecial type designed such that the trailing edges of the pads contactthe shaft when the shaft is at rest. In this way, the pads center theshaft at rest and minimize shaft runout.

A particular support structure is needed to ensure both proper supportunder load and trailing edge contact at rest. Specifically, the momentgenerated by the load acting at the trailing edge must act to turn thepad away from the shaft. A bearing designed in this way satisfies theneed for a simple inexpensive bearing which can be used to support ashaft in a gap seal assembly.

By providing a bearing which can be used with a mechanical seal, it ispossible to replace ball bearings conventionally used in connection withmechanical face seals. This allows operation at higher speeds than ispossible with ball bearings and increases the durability of themechanical face seal assembly. The bearing configuration disclosedherein can be used with any face seal assembly as a replacement for ballbearings conventionally used.

Another feature of the present invention is the possibility of aself-aligning construction which one of the rings automatically alignsitself relative to the other ring to ensure that the opposed sealingfaces are parallel to one another. The invention also provides for thepossibility of a self biased sealing ring in which a spring function isbuilt into the sealing ring to eliminate a separate spring or assist aseparate spring.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a fragmentary radial cross-section through a known gap seal.

FIG. 2 is a fragmentary front elevational view of a one known sealingmember which can be used with a gap type seal of the type shown in FIG.1 showing the spiral grooves therein; and

FIG. 3 is a fragmentary front elevational view of another known sealingmember which can be used with a gap type seal of the type shown in FIG.1 showing the T-shaped grooves formed therein.

FIG. 4 is a fragmentary front elevational view of another known sealingmember which can be used with a gap type seal of the type shown in FIG.1 showing the tapered lands formed therein.

FIG. 5 is a simplified fragmentary radial cross-section showing a gapseal which includes a sealing member according to the present invention.

FIG. 6 is a view of the top surface of a sealing member schematicallyshowing the two sections of the sealing members of the presentinvention.

FIG. 7A is a fragmentary front or top elevational view of a sealingmember according to the present invention.

FIG. 7B is a fragmentary cross-section schematically showing a portionof the sealing member of FIG. 7A in the static unloaded state.

FIG. 7C is a fragmentary cross-section schematically showing a portionof the sealing member of FIG. 7A in the loaded state.

FIG. 7D is fragmentary back or bottom elevational view of the sealingmember of FIG. 7A.

FIG. 8A is a fragmentary front or top elevational view of a sealingmember according to the present invention.

FIG. 8B is a fragmentary back or bottom elevational view of the sealingmember of FIG. 8A.

FIG. 9A is a perspective view of a sector shaped thrust pad with arrowsindicating the side lines for the top side and edge views.

FIG. 9B is a perspective view of a circular thrust pad.

FIG. 10 is a top view of a thrust pad provided with radius cuts on bothedges.

FIG. 11 is a side view of a thrust pad with tapered edges.

FIG. 12 is an edge view of a thrust pad provided with side edge rails.

FIG. 13 is a side view of a thrust bearing pad having grooves formed inthe bottom proximate the side edges.

FIG. 14A is a top view of a thrust bearing in which the individual padsare defined by pad defining grooves.

FIG. 14B is a sectional view of the thrust bearing of FIG. 14A along thelines indicated in FIG. 14A.

FIG. 15A shows a top view of a thrust bearing pad formed with a bottomrecess indicated in phantom.

FIG. 15B shows a side view of the thrust bearing pad of FIG. 15A.

FIG. 16A is a top view of a thrust bearing pad formed with a bottomrecess on each edge indicated in phantom.

FIG. 16B is a side view of the thrust bearing pad of FIG. 16A with thebottom recesses indicated in phantom.

FIG. 17A is a top view of a thrust bearing.

FIG. 17B is a bottom view of the thrust bearing of FIG. 17A.

FIG. 17C is a cross-section through the lines indicated in FIG. 17A.

FIG. 18A is a bottom view of a seal ring according to the presentinvention.

FIG. 18B is a partial sectional view of the seal ring of FIG. 18A alongthe lines indicated in FIG. 18A.

FIG. 18C is a top view of the seal ring of FIG. 18A.

FIG. 19A is a top view of another sealing ring according to the presentinvention.

FIG. 19B is a bottom view of the sealing ring of FIG. 19A.

FIG. 19C is a schematic view of the lift pads of the sealing ring ofFIG. 19A.

FIG. 20A is a top view of another sealing ring according to the presentinvention.

FIG. 20B is a bottom view of the sealing ring of FIG. 20A.

FIG. 20C is a schematic view of the lift pads of the sealing ring ofFIG. 20A.

FIG. 21A is a top view of an easily moldable thrust bearing according tothe present invention.

FIG. 21B is a bottom view of the bearing of FIG. 21A.

FIG. 21C is an exploded cross-section along the lines indicated in FIG.21A.

FIG. 21D is a bottom view illustrating modifications of the bearingillustrated in FIGS. 21A-C.

FIG. 22A is a top view of another easily moldable thrust bearingaccording to the present invention.

FIG. 22B is a bottom view of the bearing of FIG. 22A.

FIG. 22C is a partial cross-section showing the support structure forthe bearing pads in the bearing of FIGS. 22A and 22B.

FIG. 23 is a side view of a deflection pad bearing for use in the sealof the present invention.

DETAILED DESCRIPTION OF THE DRAWINGS

It is accepted that gap type seals of the type described in connectionwith FIGS. 1-4 provide good performance characteristics. The maindrawback with these designs is their cost and the difficulty ofmanufacturing and susceptibility to damage caused by wear.

The present invention begins with the proposition that it is desirablein a mechanical sealing arrangement to provide surface formations suchas wedges, grooves or recesses on the front side of either the matingring or the sealing ring so as to provide a hydrodynamic sealing effect.The present inventor has found that it is better to provide thesegrooves or recesses through deflection under load than by physicallyproviding grooves on the front surface. For instance, relatively deepgrooves formed in the backside yield the desired small deflections orrecesses or grooves on the front side without the difficulty ofmanufacturing very small grooves on the front side.

Another advantage is that the grooves on the front side resulting fromdeep grooves on the backside have a smooth sloping transition--there isno step like transition as when grooves are formed directly on thesurface. A step-like transition as in prior known constructions reducesthe hydrodynamic effect. Thus, it is desirable, as in the presentinvention, to maintain laminar flow through the pressure build up.

As noted above, the present invention is directed, in part, to animproved design of one of the two rings in a mechanical seal. Inaccordance with the present invention the improved ring can be designedas an insert to be carried on either the mating ring 10 or sealing ring13. The advantage of using an insert type ring is that the ring can bemanufactured out of a different material than the ring on which it issupported. This makes it possible, for example, to use rings formed oflow cost high performance engineering plastics which can be formed tospecification at a fraction of the cost of machining or etching metal onsilicon carbide.

FIG. 5 shows generally how a seal ring 13 can be designed to act as asupport for carrying a separate seal ring 113. According to the presentinvention the seal ring 113 can be secured to the seal ring 13 in anyknown manner including mechanical interconnection such as splines,threads, axial pins and/or by adhesion or by welding. The assembly shownin FIG. 5 is generally similar to the known assembly shown in FIG. 1.Identical reference numerals are used for similar components. The spring20 for the assembly of the present invention is, however, indicatedschematically. This is because the spring may be a separate spring as inthe prior art or it may be built into the seal ring support structure asdescribed below.

At this point it should be noted that a variety of supports arecontemplated for the seal rings of the present invention. These includeunitary supports and deflecting supports. These various supports arediscussed below. Before discussing the various supports, however, theprinciples of the sealing ring of the present invention will bediscussed. It should be understood that although specific reference ismade to a sealing ring, the ring constructions of the present inventioncan be applied to the mating ring as well.

FIG. 5 also schematically depicts several possible locations forbearings 30 to support the shaft 11. The bearings are preferablyhydrodynamic bearings according to the present invention, butconventional bearings could be used. The bearings 30 may be located atany of the positions shown or any combination or some other convenientlocation. If the bearings 30 are in contact with the seal components 10and 13 as shown, then it is possible to manufacture and sell the sealand bearings as a unit. It is even possible to manufacture thecomponents 10 or 13 as one piece with the bearings 30.

FIG. 6 shows schematically, the principle of the improved ring of thepresent invention. Specifically, the ring is divided into concentricportions 113I and 113O. One of these portions is designed to operate asa sealing portion and the other portion is designed to operate as abearing portion. The sealing portion is similar to a conventionalsealing ring; it is continuous and has a rigid support. The surface ofthe sealing portion is preferably smooth and flat. The bearing portioncomprises a series of circumferentially spaced sections which are shapedand/or supported by a support structure for deflection to achievehydrodynamic effects under load.

The relative arrangement of the sealing portion and the bearing portiondepends on the specific application. Thus, in some cases the bearingportion will be located in the radially outer portion 113O of the ring113. In other cases, the sealing portion will be in the radially outerlocation 113O and the bearing portion will be in the radially innerlocation.

One advantage of locating the bearing portion outside the sealingportion is that such an arrangement maximizes the radial distance of thepads from the center of the ring--this increases the surface velocity ofthe pads and improves film characteristics. Another factor to considerin determining whether to locate the bearing portion inside or outsideof the sealing portion is the effect of centrifugal forces on the fluidbeing sealed. In most known constructions the grooved section isprovided radially outward of the non-grooved or sealing section. Thissuggests that the bearing section should be located outside the sealingsection, i.e., in outer portion 113O. On the other hand, in theinventor's previous design, the lift pads are located inside the fluiddam. This suggests that the bearing portion should be located inside thesealing portion, i.e., in inner portion 113I.

Regardless of the relative position of the sealing portion and thebearing portion, the ring design of the present invention is flexible.The designs described below will show one arrangement or the other.

It should be noted that when the sealing portion is shown as theradially outer portion 113O and the bearing portion is shown as aradially inner portion 113I, it is possible to have the reversearrangement, i.e., the bearing portion may be the radially outer portion113O and the sealing portion may be the radially inner portion 113I.

The present invention involves shaping or supporting discrete segmentsor sections of the bearing portion so as to cause deflection of thesurface of the bearing portion under load to create a hydrodynamicpressurization of fluid which causes separation of the adjacent rings,the mating ring 10 and the sealing ring 113 at the sealing face 12, 14whereby the mechanical seal is created.

There are, of course, many different ways to shape or support thesurface of the bearing portion to create a hydrodynamic effect.

In accordance with a first illustrative embodiment of the presentinvention, the bearing portion is provided with circumferentially spacedundercuts so that the continuous surface of the bearing portion hasvariations in stiffness across its surface such that under load, thepockets or recesses are formed in those portions of the bearing portionswhich are undercut.

FIGS. 7A-7D illustrate a first seal ring construction in whichcircumferentially spaced undercuts are provided to achieve the desiredsurface formations on the top surface of the seal ring under load. Inparticular, FIG. 7A shows a section of the top surface 113T of a sealring 113 according to the present invention. As shown therein, the topsurface 113T of the seal ring 113 is completely flat. In order tooperate as a spiral type seal ring, however, it is desired that underload spiral type recesses be formed on the top surface of the seal ring113 in the positions generally indicated in phantom at 122r. If suchrecesses are provided under load, the seal ring will operate in much thesame way as a seal ring having grooves formed on the top surface bymachining or etching. In accordance with the present invention, thesedesirable surface formations are formed not by machining directly underthe top surface as in the prior art, but instead, by providingappropriate undercuts in the bottom surface 113B of the ring 113.

The controlling principles here can be best understood with reference toFIGS. 7B and 7C. FIG. 7B shows somewhat schematically a portion of theseal ring 113. As shown therein, the top surface 113T is smooth and flatin an unloaded or static situation. The bottom surface 113B is, however,provided with a groove 122 which extends into the ring 113 to undercut aportion of the top surface 113T of the ring 113.

FIG. 7C schematically illustrates the deflection of the top surface 113Twhen the top surface is subjected to a load such as by a fluid pressureat start-up of rotation of the seal ring. As shown therein, a portion ofthe top surface 113T which is undercut by the groove 122 tends todeflect downward to provide a recess 122r. Moreover, the recess 122follows the general contour of the undercut 122.

Another important advantage of the present invention is that the recesshas a gradual taper to it so that there is no sharp step or drop. Thisincreases the likelihood of laminar flow across the top surface 113T.Thus, it can be seen that the desired recesses shown in FIG. 7A at 122rcan be formed under load by providing undercuts in the fashionillustrated at 122 in FIG. 7D.

It should be appreciated that it is much easier to provide therelatively deep grooves 122 in the bottom surface 113B of the seal ringthat it is to provide extremely fine grooves in the top surface 113T ofthe seal ring which generally must be perfectly smooth.

FIGS. 8A and 8B illustrate another embodiment of the present invention.In this embodiment, the seal ring 113 is once again smooth and flat onthe top surface 113T as shown in FIG. 8A. In this case, however, it isdesired to form a cirfumferentially spaced series of T-shapeddepressions on the top surface 113T under load. The desired depressionsare again indicated in phantom at 122r. In order to provide thedepressions 122r desired, the bottom surface 113B of the seal ring isundercut in the manner shown in FIG. 8A. The contour of the undercuts asshown at 122. In the same manner as described above, provision of theundercuts 122 causes deflection of the top surface 113T under load.

Once again, the construction of the present invention offers numerousadvantages to machining or etching the recesses directly on the topsurface. Among these is the relative ease of providing deep grooves inthe back surfaces opposed to the difficulty of providing microfinegrooves on the top surface. Moreover, the transition between therecesses and the non-recessed portions of the top surface 113T is muchmore smooth when the recesses are achieved through deflection byundercutting the top surface.

In accordance with another construction according to the presentinvention, the bearing section may be shaped and/or provided with asupport structure which essentially divides the bearing section into acircumferentially spaced series of discrete bearing or lift pad portionswhich are connected to the sealing portion. The lift pads are shapedand/or supported so as to cause hydrodynamic deflection under load. Thelift pads essentially operate so that the friction forces which aredeveloped will tend to rock the pad so that the leading edge of the padface will move away from the mating ring upon rotation of the shaft. Awedge of fluid will develop between the mating ring and the pad face ofthe pad. Further, by providing an appropriate support structure asdiscussed below, the pads may also move to compensate for shaftmisalignment and equalize loading among pads.

The most important consideration in the design of lift pads is the shapeof the space, typically a converging wedge, between the lift pad and thesurface it opposes. Since the shape of the surface the lift pad opposesis basically invariable, it follows that the most importantconsideration in the design of hydrodynamic lift pads is the shape ofthe pad surface under load. The shape of the pad surface under loadprincipally depends upon two factors: the shape of the pad itself andthe construction and location of the pad support structure.

For purposes of this description, the various pad designs will bediscussed first followed by a discussion of various support structuredesigns. It must be emphasized that the various support structuresdisclosed herein can be used with any of the pad shapes disclosed hereinand the pad-shapes used herein can be used with any of the supportstructures disclosed herein.

The pads and support structure are designed to optimize the shape of theconverging wedge formed between the pad surface and the shaft when theshaft rotates. As detailed in applicant's pending application Ser. No.07/685,148 which has been incorporated herein by reference, this can bedone by modifying the pad shape, the support structure or both.Specifically, the pad can be modified to include grooves, cuts, railsand recesses to achieve desired deformations under load. The supportstructure can be designed to support the pads for movement in the sixdegrees of freedom (i.e., translation or movement in the +x, -x, +y, -y,+z and -z directions) and rotation about the x, y, and z axes so as tooptimize formation of the hydrodynamic wedge.

The lift pads and support structure used in the seal rings of thepresent invention may be designed in three dimensions to providedeflection with six degrees of freedom so as to ensure optimum wedgeformation at all times.

In computer analysis of this system using a finite element model, theentire lift pad can be treated as a completely flexible member thatchanges shape under operating loads. By adding more or less flexibilitythrough design of the basic structure, lift pad characteristics may beachieved that provide stable operation over wide operating ranges.

The present invention allows for movement of the lift pad in anydirection (i.e., six degrees of freedom) to form a converging wedgeshape; allows for the pad itself to change shape (e.g., flatten) toimprove performance; and allows the lift pads to compensate formisalignment of the supported part or shaft and to equalize loadingamong the lift pads.

The present inventor has discovered that important performancecharacteristics can be achieved by simply modifying the lift pad shape.Consequently, the support structure can be simplified, and in somecases, even eliminated.

Examples of typical lift pad shapes according to the present inventionare illustrated in FIGS. 9A and 9B. FIG. 9A shows a sector shaped pad132. The sight lines for a top view T, an edge view E and a side view Sare indicated by arrows labeled T, E and S, respectively. FIG. 9B showsa circular pad 20. The arrows indicate the sight lines for the top viewT, edge view E and side view S discussed below. These pad shapes are allcharacterized by uninterrupted planar surfaces and a uniform padthickness.

Various modifications to these pad shapes will be discussed hereinafter.It should be kept in mind that any of these modifications to the shapeof the pad may be used in combination or alone. Also, the modificationscan be easily adapted to pads having shapes other than the specific padshapes illustrated. Moreover, the pads may be symmetrically shaped toallow bidirectional operation or non-symmetrically shaped to providedifferent operating conditions depending on the direction of rotation.The modified pad shapes discussed hereinafter may be used in combinationwith any support structures including those described in thisapplication where appropriate or, when used in the proper combination,may eliminate the need for a deflecting support structure altogether.

The first possible modification to the general pad shape is shown inFIG. 10.

This modification is based on finite element analysis which has shownthat, in some instances, increasing the length of the edge where thelubricant enters (the leading edge) allows more lubricant to be directedtoward the pad center. To achieve this effect, a radius cut may beformed on the pad surface to lengthen the leading edge. The cut may beformed either entirely through the pad or partially through the padsurface to provide a recess in the pad surface. It should be kept inmind that the provision of such a radius cut decreases the load bearingsurface of the pad. Thus, there is a trade off; more lubricant but lessload bearing surface.

FIG. 10 shows a top view of a lift pad 132 in which a radius cut 132C isformed as shown. In the illustrated embodiment, the cut 132C is providedon each edge of the pad 132. This is because the illustrated pad isintended for bidirectional use and the improved result is desired inboth directions. If unidirectional operation is sufficient, the cutshould only be provided on one edge.

FIG. 11 illustrates another possible modification to the basic padshape. Specifically, it has been learned that tapering the leading edgeof the lift pad results in increased inlet bending. This allows morelubricant to enter into the space between the lift pad and the surfaceit supports thus increasing the load carrying capability of the pad.Finite element analysis using computers can predict the amount ofbending needed to obtain optimum lubricant flow.

FIG. 11 is a side view along the S axis in FIG. 9A illustrating a liftpad 132 with a taper 132t formed at each edge. The taper is provided ateach end to allow for bidirectional operation. Of course, ifunidirectional operation is sufficient, only one edge, the leading edge,should be tapered.

The basic pad shape may also be modified by providing rails on the sideedges of the pads such that, under load, the pad deflects to form achannel which retains lubricant on the pad face and minimizes end orside leakage.

FIG. 12 shows an edge view of a lift pad 132 provided with side edgerails 132r on the radially inner and outer edges. The deflection of thispad under load (greatly exaggerated) is indicated in phantom. As can beseen, the pad deflects under load to form a lubricant retaining channel.

As mentioned before, it is sometimes desirable to increase the inletbending of the leading edge of a lift pad. Another modified lift padshape for achieving or enhancing this desired result is shown in FIG.13. This drawing shows that in addition to or instead of tapering theleading edge, a groove may be formed on the lower edge of the lower sideof the pad proximate the leading edge to cause increased leading edgebending while maintaining a flatter surface. FIG. 13 shows a lift pad132 with grooves 132g formed in the bottom near both edges to allowbidirectional operation.

Another consideration in the design of hydrodynamic lift pads is thatthe pads themselves may be formed from a single member by simplyproviding grooves to define individual pads. FIGS. 14A-14B show how acontinuous surface can be divided into individual lift pads 132 throughthe provision of pad defining grooves 132p. In this case, FIG. 14A is atop view and FIG. 14B is a side view along the lines indicated in FIG.14A. In these drawings, the continuous sealing portion is notillustrated. As discussed below, a sealing portion may be connected tothe illustrated bearing portion by thin ligaments.

A final consideration in the design of specific lift pad shapes is theprovision of bottom recesses on the pads. Specifically, the provision ofbottom recesses can cause channeling in a manner somewhat like thatshown in FIG. 12 and allow inlet bending in a way such as the taperedstructure shown in FIG. 11. FIGS. 15A-15B show top and side views of alift pad 132 formed with a bottom recess 132b to cause channeling. Thereduced pad area also enables compressive deflections onto the bottomsurface which develops a converging wedge. Since this modification isprovided on only one edge of the pad 132, the pad is intended for use ina unidirectional seal.

FIGS. 16A-16B illustrate lift pad configurations similar to those shownin FIGS. 15A-15B except that the bottom recesses 132b are provided atboth edges of the lift pad so as to permit bidirectional operation.Specifically, the lift pad 132 shown in FIGS. 16A and 16B includesbottom recesses 132b at each edge thereof. As is apparent by comparingFIGS. 16A-16B with FIGS. 15A-15B, the bottom recesses are somewhatsmaller to accommodate the provision of such recesses at each edge.

As mentioned before, the design of a lift pad shape for any particularapplication depends on the requirements of that application. Theforegoing structural modifications and considerations can be used aloneor in combination. FIGS. 17A-17C show how all these features can becombined in a single bearing portion. This is not to suggest that allthese features should necessarily be included in every design. Indeed,this would rarely be required. However, it is possible to combine allthese features in a single bearing portion, if desired.

FIG. 17C shows a top view of a bearing portion in which the lift pads132 are provided with radius cuts 132C to increase the length of theleading edge. The radius cuts 132C are provided on each edge so as topermit bidirectional operation. Of course, if desired, the radius cuts132C can be provided on only one edge to provide optimum results forunidirectional operation. In FIG. 17C, the individual pads 132 aredefined by pad defining grooves 132p.

FIG. 17B shows a bottom view of the bearing portion of FIG. 17A. In thisview, it can be seen that the lift pads include bottom rails 132r,grooves 132g to increase inlet bending, a taper 132t to further increaseinlet bending and a bottom recess 132r to further channel liquid andincrease inlet bending. In this case, the grooves, taper, and recessescollectively provide the desired deflection.

FIG. 17C is a cross-section of the bearing portion of FIG. 17A along thelines indicated in FIG. 17A. FIG. 17C also shows that the bearingportion mounted includes a support structure 137. The support structureis shown schematically as a box to indicate that, in accordance with thepresent invention, any of the support structures disclosed herein can beused. As noted above, it is possible through proper pad design toobviate the need for a deflecting support structure. In such a case, thesupport structure could be rigid, e.g., the housing. Alternatively,however, the support structure can be a deflecting support structure ofany of the types disclosed herein having primary, secondary and tertiarysupport portions for supporting the lift pads for movement with sixdegrees of freedom. Likewise, the pad modifications discussed herein aregenerally applicable individually or in combination to the lift pads ofany of the bearing portions disclosed herein.

Another complete seal ring design is depicted in FIGS. 18A-18C. Thesedrawings depict a washer-like seal ring in which the desired pad shapeis obtained in an extremely thin washer-like cylindrical element. Theseal ring includes a bearing portion and a continuous flat smoothsealing portion. In this case, the sealing portion is located radiallyoutward of the bearing portion. Of course, the reverse is possible, ifpreferred. FIG. 18A is a bottom view of the washer-like seal ringshowing that circumferentially spaced bottom recesses 132b are providedon the bottom of the bearing portion. Additionally, arcuate groovesseparate the bearing portion from the sealing portion except that a thinligament connects each discrete lift pad to the sealing portion.Finally, FIG. 18C shows a top view of the washer-like seal ring showingthat circumferentially spaced pad defining grooves 132p are formed inthe top surface. The bottom recesses 132b are shown in phantom. Thegrooves in the top surface together with the bottom recesses define aplurality of discrete circumferentially spaced lift pads 132. Thecross-section of the grooves 132p and bottom recesses 132b is bestillustrated in the cross-sectional view of FIG. 18B. As shown in thisfigure, the grooves 132p and 132b are very shallow.

As previously discussed, the pad defining grooves 132p define acircumferentially spaced series of lift pads 132. The bottom recesses132b undercut the pad surface to a sufficient extent that the portion ofthe lift pad surface that is undercut can deflect slightly downward soas to form a converging wedge and a lubricant retaining channel.Collectively, these deflections result in the formation of a series ofconverging wedges so that a layer of pressurized fluid film is formedbetween the lift pad and mating ring. Further, because of the nature ofthe bottom recesses 132b the lubricant is retained on the pad surfaceand does not escape from the radially inner and outer edges of the pad.Finite element analysis has shown that, under sufficient load, thissimple seal ring will deflect so as to operate as a series ofhydrodynamic lift pads even without a deflecting support structure.Thus, a simple washer-type seal ring configuration of the type shown inFIGS. 18A-18C can be mounted on a rigid support structure and stillobtain satisfactory results. Of course, a deflecting support structurecould be used, if desired.

FIGS. 19A-19C show another seal ring according to the present invention.Like the previous embodiments, this seal ring includes a sealing portionand a bearing portion. In this case, the bearing portion is disposedradially inward of the sealing portion.

The bearing portion comprises a plurality of circumferentially spacedlift pads 132. The lift pads 132 are separated from one another byradially extending spaces. The pads 132 are largely separated from thesealing portion by circumferential spaces. The pads are, however,connected to the sealing portion by ligaments 132L. The ligaments aredesigned to connect the sealing portion to the lift pads so that thesealing portion is carried axially with the lift pads as a result of thepressurized fluid film between the lift pads and the mating ring. On theother hand, the ligaments are sufficiently flexible that they do notinterfere too much with the deflection of the lift pads needed to causeformation of the pressurized fluid film between the seal ring and themating ring.

In the embodiment shown in FIGS. 19A-19C, the lift pads 132 aresupported on a rigid support structure in the form of a post 137. Thepost 137 is located nearer to the outer periphery of the lift pad 132than to the inner periphery. Additionally, the lift pads tend to deflectdownward to limit centrifugal leakage.

The lift pads can be supported for deflection so as to retain thehydrodynamic fluid, thus obviating the problem of fluid leakage. The padis supported so as to tilt toward the seal ring's inner diameter underload so as to prevent centrifugal leakage. Generally, this is achievedwhen the pad support surface at which the primary support structuresupports the lift pad is located closer to the lift pad outer diameterthan to the lift pad inner diameter. When the primary support structureincludes two or more radially spaced beams, the overall supportstructure must be designed to cause deflection of the bearing pad at theinner end. Further, when the lift pad is supported by a plurality ofradially spaced beams and the region between the beams is not directlysupported, the pad will tend to deflect so as to form a concave fluidretaining channel.

The configuration of the lift pads is best shown in FIGS. 19B and 19C.As shown therein the lift pads 132 have a generally arcuate shape. Bothradially extending edges of the pads are tapered as shown at 132t asdiscussed earlier with respect to FIG. 11. The provision of such tapersallows increased inlet bending regardless of the direction of relativemovement between the lift pad 132 and the surface it opposes. Finiteelement analysis has shown that the inlet bending allowed by this liftpad configuration provides proper wedge formation under certain loadconditions.

As noted earlier, the seal rings of the present invention all include asealing portion and a bearing portion. The bearing portion may belocated either radially inward or radially outward of the sealingportion as appropriate. To illustrate this point, FIGS. 20A-20Cillustrate the seal ring configuration of FIGS. 19A-19C modified suchthat the lift pads 132 are located radially outward of the sealingportion.

The seal ring shown in FIGS. 20A-20C is similar to that of FIGS. 19A-19Cexcept for the radial arrangement of the lift pads relative to thesealing portion. In particular, the lift pads 132 are connected to thesealing portion by ligaments 132L. The lift pads 132 also include arigid post-like support structure 137 and tapered edges 132 for thereasons discussed above in regard to the embodiments of FIG. 11 andFIGS. 19A-19C.

In the above examples, the bearing pads are formed with a tapered padshape and are supported on a solid pedestal to provide a simplehydrodynamic bearing effect. If desired, the support structure for eachof the discrete bearing or lift pads sections could include any of thesupport structure configurations known from applicant's previous patentapplication. However, it is important to keep in mind that manufacturingsimplicity is an important feature of the present invention. Hence, thesimple support structure and pad configuration shown herein appears toprovide favorable results.

One manufacturing consideration is ease of molding. The seal ringconstructions of the present invention are capable of being molded bysome molding technique. However, only certain shapes can be injectionmolded in a simple two-piece mold, i.e., a mold which does not includecams. Another advantage of the seal rings of the present invention isthat they can be constructed with easily moldable shapes which aredefined as shapes which can be injection molded using a simple two-piecemold. An easily moldable shape generally is characterized by the absenceof "hidden" cavities which require cams for molding.

Other easily moldable seal rings are possible. Several suchconstructions are discussed below with reference to FIGS. 21A-22C. Indescribing these constructions only the bearing portion of the sealingring is depicted. It is to be understood that the sealing portion is acontinuous ring as in the previous embodiments and that the sealingportion is connected to the bearing portion by ligaments as describedabove. Moreover, the sealing portion may be located either radiallyinward or radially outward of the bearing portion.

FIGS. 21A-21C illustrate another easily moldable seal ring. The bearingportion includes a plurality of circumferentially spaced lift pads 132mand a support structure supporting each of the lift pads 132m. Thesupport structure includes a primary support portion which includescircumferential beams 134mb and 134ma, a secondary support portion whichincludes radially extending beam 136m and a tertiary support portionwhich includes the stub-like pair of beams 138m. It should be notedthat, in FIGS. 21A-21C, the dimensions of the support structure aresomewhat distorted to provide clarity. For instance, as shown in FIG.21C, the circumferential beams 134ma and 134mb are shown as extremelythick. Such a beam structure would provide a very rigid support for thelift pads 132m and in practice, such a rigid support would probably notbe necessary or desirable.

Variants of the specific moldable beam structure illustrated arepossible. For instance, either or both of the spaced circumferentialbeam segments 134ma or 134mb could be formed as a continuouscircumferential beam element. Additionally, the secondary supportportion could include a plurality of radially extending beams betweeneach lift pad 132m. Further, the primary support structure could bemodified to include three or more circumferential beam segmentsconnecting each pair of adjacent lift pads and/or circumferential beamsegments; also, segments of different radial widths could be used.Further, the stub-like beam portions 138m could be provided along theradially extending edges of the beams 136 rather than thecircumferentially extending ends. Finally, as with any seal ring supportstructure in accordance with the present invention, the structure couldalso be varied by varying the length or thickness of any of the elementsin the support structure to modify the deflection characteristics of thesupport structure.

In order to illustrate a number of possible support structureconstructions, FIG. 21D depicts a different support structure for eachof the lift pads 321m-326m. In particular, FIG. 21D is a bottom viewwith the modifications illustrated herein. It should be understood thatthese various support structures are shown in a single seal ring forpurposes of illustrating the present invention. In normal use, each ofthe lift pads 321-326m would have a similar, though not necessarilyidentical, support structure to assure uniform performance.

The support for lift pad 322m differs from that for the lift pads 132min that an oval shaped projection extends from the back of the bearingpad surface to provide a rigid support for the outer circumferentialedge of the lift pad 321m. By virtue of this construction, the lift pad321m would be extremely rigid at its outer circumferential end.

The support for lift pad 322m is similar of that to 321m except thatrather than a single large projection, two smaller projections 122mextend from the bottom of the bearing portion proximate the outercircumferential edge of the lift pad. Like the projection 120m, thesetwo projections 122m provide rigidity to the outer circumferential edgeof the lift pad 322m. However, this construction allows the lift pad todeflect in the unsupported region between the projections.

The lift pad 323m is supported by a modified support structure whichincludes a continuous circumferential beam 134ma in the primary supportportion. Similarly, the lift pad 324m includes a continuous innercircumferential beam 134mb. The provision of such continuous beamsincreases the rigidity of the bearing support structure.

The support structure for lift pad 325 is modified by the provision oflarge openings 142m in the inner beam 134mb and smaller openings 144 inthe outer beam 134ma. The provisions of these openings increase theflexibility of the beams. Naturally, the larger openings increase theflexibility of the beams to a greater extent than the small openings144. Variants of this support structure include the use of differentsized openings or a different number of openings to bias the lift pad325m in a predetermined direction.

The lift pad 326m is supported by a modified structure in which theprimary support portion includes a membrane 134m rather than a pair ofbeams. In the illustrated example, one of the membranes is provided witha opening 146 to bias the lift pad 326m in a predetermined direction. Ofcourse, the provision of the opening 146m is not necessary and ifdesired, a number of openings could be provided.

As is evident from these drawings, the moldable seal rings do notinclude any hidden cavities which would necessitate the use of a complexmold and/or a mold including a displaceable cam. In particular, sinceeach surface of the seal ring structure is directly visible in eitherthe top view of FIG. 21A or the bottom view of FIG. 21B, the bearing canbe simply molded using a two piece mold. Specifically, a first moldpiece defines those surfaces which are directly visible only in the topview of FIG. 21A. The second mold piece defines those surfaces which areonly visible in the bottom view of FIG. 21B. Surfaces having edgesvisible in both FIGS. 21A and 21B can be molded using either or bothmolds. In the illustrated seal ring, easy moldability is achievedbecause the secondary and tertiary support portions arecircumferentially located in the space between lift pads. Themodifications illustrated in FIG. 21D do not alter the easy moldabilityof the seal ring.

More complex variants of the moldable seal ring illustrated in FIGS.21A-21D are possible. In particular, any of the previously discussedmodifications of the seal ring structure which can be adapted to easymolding could be employed. For instance, the primary support beams couldbe continuous. Thus, the provision of an easily moldable seal ring doesnot necessarily require a simple seal ring construction. An example of amore complex seal ring structure is illustrated in FIGS. 39A-39C.

As illustrated in FIGS. 22A-C, the seal ring includes a plurality ofcircumferentially spaced lift pads 232m supported by a lift pad supportstructure. The secondary and tertiary portions of the support structureare similar to corresponding portions of the support structure of FIGS.21A-21D. However, the seal ring of FIGS. 22A-C differs from the sealring of FIGS. 21A-D in that, in the seal ring of FIGS. 22A-C, theprimary support portion includes a plurality of complex beams 234.Specifically, each lift pad is supported by a radially outer continuouscomplex circumferential beam 234ma. The pads are further supported bythe plurality of spaced circumferential complex beams 234mb. The complexshapes of the continuous beam 234ma and the beam segments 234mb can bebest appreciated with reference to FIG. 39C which shows, somewhatschematically, the profile of the complex beams 234. In operation, thebeams 234ma and 234mb function as a beam network.

Thus, it can be seen that numerous complex seal ring constructions canbe provided while retaining the ability to mold the seal ring with asimple two-piece mold, i.e., easy moldability. Naturally, each structureprovides unique deflection characteristics which must be considered indesigning the seal ring for optimum wedge formation.

The present invention also relates to the structure for maintainingstability of the sealing ring and the mating ring and to maintainingparallel alignment of their faces. In the absence of stability andparallelity, the rings may distort clockwise or counterclockwise due toexcessive heat generation or heat removal resulting in contact betweenfaces with subsequent face damage or seal destruction. In accordancewith another aspect of the invention, these pressure and temperaturedeflections are minimized by the self-aligning support structureprovided on at least one of the sealing ring and the mating ring.

In the preferred embodiment of my invention, this self-aligning featureis obtained by providing a flexible support structure for the sealingring.

A unique aspect of the present invention is that in those seal ringconstructions in which the lift pads are supported by beam-likestructures, the necessary spring force can be built into the beam-likesupport structure. This is because such a support structure can supportthe lift pads for movement with six degrees of freedom. Through designof the support structures in terms of thickness and orientation of thebeams, the appropriate spring constant can be provided to the lift pads.By way of example with reference to the embodiments of FIGS. 22A-22C itcan be seen that the lift pads 232 are supported on a beam-like supportnetwork best shown in FIG. 22C so that the pads are supported fordeflection up and down as shown in FIG. 22C. The resistance to suchdeflection, i.e., the spring constant of the support structure, isdetermined by the rigidity of the beam network which supports the liftpads. In effect, the beam network operates as a spring. Thus, it can beappreciated that the appropriate design of the support structure for thelift pads and the need for a separate spring element can be eliminated.

A beam-like support structures of the type shown in FIGS. 22A-C can beused as a separate support for seal rings which do not have flexiblesupport structures. Thus, for example, in the case of a seal ringconstruction such as that shown in FIGS. 7-8, the seal ring itself mightbe supported on a beam-like spring support structure instead of a moreconventional spring of the type shown in FIG. 5. It is easier to use abeam-type spring network in those cases where the support structure isflexible since this eliminates the need for a separate spring.

Understanding that the support structure can be used to provide thenecessary spring force for a gap-type seal assembly, it can be easilyappreciated that the support structure itself ensures proper alignmentof the sealing face with respect to the face of the mating ring so as toavoid the problems discussed above resulting from instability. Thus, inaccordance with the present invention, the beam-like support structuresdescribed herein can be used to provide a self-aligning construction inwhich one of the rings automatically aligns itself relative to the otherring to ensure that the opposed sealing faces are parallel to oneanother. This construction also provides a self-biased sealing ring inwhich a spring constant is built into the seal ring support structure toeliminate a separate spring or assist a separate spring.

In accordance with the present invention, the need for close tolerancesbetween the bearing pad and the shaft portion to be supported can beobviated by dimensioning the bearing so as to eliminate the spacingbetween the bearing pad and the shaft portion to be supported.

As noted previously, there is still a need for a hydrodynamic radialbearing which can be used in connection with a gap-type seal.Conventionally, only rotating element type bearings are suitable forgap-type mechanical seals because of the risk of mechanical run out.Specifically, in any mechanical gap-type seal, it is essential that theshaft remain centered. In other words, the shaft should not be allowedto float within a radial envelope as is conventional in hydrodynamicbearings. In rotating element type bearings, shaft centering is not aproblem because the shaft is in effect maintained in solid contact withthe housing. With conventional hydrodynamic bearings, however, the shaftis separated from the housing by a spacing known as the radial envelopeand in operation the shaft is supported on a fluid film. Thus, becauseof the spacing between the shaft and the bearing surface in conventionalhydrodynamic bearings, the center of the shaft tends to vary duringoperation. This movement of the shaft leads to a problem known as "shaftrun out" which defeats the operation of the mechanical seal.

The improved gap seal of the present invention also contemplates the useof an improved deflection pad bearing that provides better shaftcentering over the operating speed range than current bearings whilealso maintaining good rotor dynamic stability. Specifically, the presentinvention provides a new bearing design in which the stabilitycharacteristics of traditional bearings are maintained while alsoproviding better shaft centering characteristics.

Specifically, the mechanical gap seal of the present inventionpreferably includes a deflecting tilt pad bearing designed to maintainshaft centering. The bearing construction is a deflection pad bearing ofthe type described in applicant's previous application Ser. No.07/685,148 which has been incorporated herein by reference. However, thebearing is somewhat different in that at least some of the bearing padsare provided with a support structure that allows the bearing pads tocontact the shaft at rest. These bearing pads which are in contact withthe shaft at rest maintain the shaft in the proper centered position.

FIG. 23 shows a side view of one hydrodynamic bearing suitable for usein the mechanical gap-type face seal of the present invention. In themanner described in the previous application incorporated herein byreference, this bearing may be formed by cutting cuts and grooves in acylindrical bore. The bearing can, of course, also be manufacturedaccording to any known technique.

FIG. 23 shows a cross-section of the three pad positive-centeringdeflection pad bearing 30 developed for use with the mechanical seal ofthe present invention. For ease of description the bearing will bedescribed as if it was formed from a cylindrical bore. As shown in FIG.23, the bearing includes two different types of bearing pads. The firsttype is relatively short pads indicated at 32. The second type isrelatively long pads indicated at 31. The relatively short pads 32operate as flexible centering pads. These pads begin at zero clearancei.e., shaft contact and expand at design speeds under hydrodynamicpressure to a larger operating clearance. The larger, centrally pivotedpads 31 offer low pivot rotational stiffness and high radial stiffnessto attenuate unbalanced response and maintain stability.

The operation of the larger centrally pivoted pads 31 is similar toconventional pivot type pads for use in hydrodynamic bearings. Inparticular, the pads 31 are supported by a single ligament 37 for simpleflexibility in the plane of the paper of the drawing. In essence, thelarger pads 31 have a support structure 37 which allows simple pivotingof the pad 31. It is known that a simple pivoting construction like thisprovides adequate support at operational speeds.

To ensure proper shaft centering and support at start up, the smallerpads 32 have a more flexible support structure. Specifically, thesupport structure of the pads 32 includes a primary support portion 371in the form of a stub type shaft, a secondary support portion in theform of a elongated circumferential beam 372 and a tertiary supportportion 373 in the form of a stub shaft connecting the secondary supportportion 372 to the base or outer periphery of the bearing 30. Because ofthe elongated nature of the beam 372, any force acting on the surface ofthe pad 332 causes pivoting about a pivot point PP as is clearly evidentin FIG. 23. This pivot point PP is located circumferentially beyond thetrailing edge 32t of the bearing pads 32. As a consequence, any forceacting on the pad 32 causes a moment M to be generated in the directionshown in FIG. 23. This ensures that under static state, contact betweenthe pad 32 and the shaft occurs only at the trailing edge of the pad. Inthis way, a preformed wedge is formed even in the static state.

Moreover, the contact between the trailing edge of the pad 32 and theshaft maintains the shaft center in the proper position. Thus,immediately upon start up a hydrodynamic effect occurs while at the sametime the shaft is properly centered and shaft run out does not occur.Thus, this hydrodynamic bearing, unlike known hydrodynamic bearings, canbe used in a mechanical face seal of the gap type. The advantages ofhydrodynamic bearings versus rolling element type bearings are set forthin applicant's previous application which has been incorporated hereinby reference. An additional important advantage occurs at high speedwhere rolling element bearings are subject to rapid wear, buthydrodynamic bearings perform without wear because there is noshaft-to-pad contact.

The particular bearing construction shown in FIG. 23 includes two typesof pads support structures. This provides a wide range of supportconditions. It should be understood, however, that bearings could bedesigned such that each of the pads are supported in the same way. Ifsuch a bearing is to be used with a mechanical face seal of the gaptype, then each support structure should, like the support structure ofthe bearing pads 32 be designed such that the pivoting caused by staticload occurs beyond the trailing edge of the pad in the manner shown inFIG. 23 so that under static loading the pads are deflected so that onlytheir trailing edge contacts the shaft. This ensures that a preformedwedge is formed and at the same time properly centers the shaft.

What is claimed is:
 1. In a gap type rotary mechanical seal assembly, amethod of providing a high pressure fluid barrier between a sealingmember and a rotating surface to be sealed comprising the steps of:securing the sealing member to a bearing member; supporting the bearingmember for deformation under load relative to the surface to be sealed;providing a supply of compressible fluid in the proximity of the bearingassembly; securing the surface to be sealed to a shaft for rotation withrespect to the bearing member so as to cause the bearing to deflect in away that creates a pressurized fluid film between the bearing and thesurface to be sealed so as to axially move the sealing member secured tothe bearing member away from the surface to be sealed whereby a barrierof high pressure fluid is formed between the sealing member and thesurface to be sealed.
 2. A one-piece seal ring for use in a mechanicalface seal; the seal ring comprising a continuous sealing portion and aplurality of discrete lift pads integrally formed with the continuoussealing portion, the lift pads being supported for deflection under loadso as to create a pressurized fluid film.
 3. The one-piece seal ring ofclaim 2, wherein the lift pads are supported by a support structure thatis integrally formed with the lift pads.
 4. The one-piece seal ring ofclaim 3, wherein the support structure is a single post extending fromthe lift pad.
 5. The one-piece seal ring of claim 2, wherein the liftpads comprise circumferentially spaced edges that are tapered.
 6. Theone-piece seal ring of claim 2, wherein the lift pads are locatedradially inward of the continuous sealing portion.
 7. The one-piece sealring of claim 2, wherein the lift pads are separated from one another bycircumferentially extending gaps.
 8. The one-piece seal ring of claim 2,wherein the lift pads are connected to the continuous sealing portion byradially extending ligaments.
 9. The one-piece seal ring of claim 2,wherein the lift pads each include an arcuate radially inner edge, anarcuate radially outer edge and two spaced radially extending edges thatextend between the inner and outer edges.
 10. The one-piece seal ring ofclaim 2, wherein at least one of the radially extending edges of each ofthe lift pads is tapered.
 11. A seal ring for use in a mechanical faceseal, the sealing ring comprising: a continuous sealing portion having aradially inner edge and a radially outer edge; and a plurality of liftpads, each of the lift pads having an edge that is connected to an edgeof the continuous sealing portion.
 12. The seal ring of claim 11,wherein a ligament connects an edge of each of the lift pads to an edgeof the continuous sealing portion.
 13. The seal ring of claim 12,wherein each of the lift pads includes a radially outer edge, a radiallyinner edge and two spaced radially extending edges that extend betweenthe radially outer edge and the radially inner edge.
 14. The seal ringof claim 13, wherein a ligament extends between the radially inner edgeof the continuous sealing portion and the radially outer edge of each ofthe lift pads to provide the connection between the continuous sealingring and the plurality of lift pads.
 15. The seal ring of claim 14,wherein the continuous sealing portion, the ligaments and the lift padsare all integrally formed as a single piece.
 16. The seal ring of claim11, wherein the lift pads are supported for deflection under load. 17.The seal ring of claim 16, wherein the lift pads are supported by asupport portion formed integrally therewith.
 18. A seal ring for use ina mechanical face seal, the sealing ring comprising a continuous sealingportion, a bearing portion spaced radially from the continuous sealingportion and a plurality of radially extending ligaments connecting thecontinuous sealing portion to the bearing portion.
 19. The seal ring ofclaim 18, wherein the bearing portion comprises a plurality ofcircumferentially spaced lift pads.
 20. The seal ring of claim 19,wherein the lift pads are separated from one another by a radiallyextending gap and each lift pad is connected to the continuous bearingportion by a radially extending ligament.